Seal devices have been used in a variety of applications to prevent fluid from leaking between joined pieces. For example, a seal device is interposed and compressed between flanged end-connections of a flow line wherein in-line process control equipment is installed. In-line process control equipment includes valves, pumps, flow meters, temperature controllers, pressure controllers and the like. This equipment usually cannot be welded into the flow line because time-scheduled maintenance requires temporary removal of this equipment and, occasionally, depleted equipment must be removed for replacement. In-line process control equipment is used in a variety of industries such as the chemical industry for processing, transporting and dispensing a myriad of chemicals and chemical compounds as well as in the oil and gas industry for recovering, distributing and processing oil, gas and by-products thereof.
Different fluid carrying pipeline systems operate at different pressures. When the operating pressure is in excess of 1000 PSI, it is often referred to as a "high pressure system". As such, the pipeline system requires high pressure flanged joints. Naturally, flanges that form a flange joint and the gaskets and seals associated therewith must meet certain pressure rating specifications, depending upon the pressure of the pipeline system with which they are intended to operate.
Accordingly, various organizations have classified flanges, gaskets and seals to certain rated pressures. For example, the American Society of Mechanical Engineers "ASME" has specified certain rules for bolted flange connections. One such example is found in the ASME Boiler and Pressure Vessel Code, Section 8, Division 2, Appendix 3. ASME/ANSI B16.5 specifications set forth certain pressure classes for flanges and their rated working pressures, as exemplified in the following table:
TABLE I ______________________________________ ASME/ANSI ASME/ANSI B16.5 Pressure Class Rated Working Pressure ______________________________________ 300 750 PSI 600 1440 PSI 900 2250 PSI 1500 3750 PSI 2500 6250 PSI ______________________________________
These working pressures are a result of the interplay of many variables included in the flange joint connection. In the design of a flange joint, two primary design conditions are typically considered. First, the conditions required to resist a hydrostatic end force tending to part the flange joint may be referred to as the "operating condition". Here, sufficient compression must be maintained on a gasket contact surface sufficient to assure a tight joint. The minimum load on the flange joint under operating conditions is a function of the design pressure, the gasket material and the effective gasket contact area to be kept tight under pressure. This formula may be expressed as follows: EQU W.sub.m1 =H+H.sub.p =0.785G.sup.2 P+(2b.times.3.14GmP) (1)
where G=diameter at location of gasket load reaction PA1 P=design pressure PA1 h=effective gasket contact surface seating width PA1 m=gasket factor PA1 where b=effective gasket contact surface seating width PA1 G=diameter at location of gasket load reaction PA1 y=gasket contact surface unit seating load PA1 where f=hub stress correction factor PA1 L=factor (te=1/T+t.sup.3 /d) PA1 g.sub.1 =thickness of hub at back of flange PA1 B=inside diameter of flange PA1 t=flange thickness PA1 e=factor (F/h for integral flanges) PA1 Y=factor involving K PA1 Z=factor involving K PA1 where W=flange design bolt load PA1 C=bolt circle diameter PA1 G=diameter at location of gasket load reaction
The conditions existing when the gasket contact surfaces seated by applying an initial load with the flange bolts is referred as "W.sub.m2 ", that is the minimum required bolt load for gasket seating. This factor is a function of the gasket material and effective gasket contact area to be seated. This relationship may be expressed as follows: EQU W.sub.m2 =3.14 bGy (2)
In addition to the bolt loading considerations of Formulas 1 and 2, other stresses on the flange joint are also significant. Among these are longitudinal hub stress (tension), radial flange stress (hoop stress) and tangential flange stress. The moment of a load acting on the flange, of course, is a product of the load (force) and its moment arm. For operating conditions, the total flange moment "M.sub.o " is the sum of three individual moments "M.sub.D ", "M.sub.T " and "M.sub.G ". For gasket seating, the total flange moment "M.sub.o " is based on flange design bolt load which is opposed by the gasket load.
In calculating flange stresses, it is important to determine such stresses for both operating conditions and gasket seating. For integral type flanges, these stresses are expressed as follows: EQU Hub stress: S.sub.H =fM.sub.o /Lg.sub.1.sup.2 B (3) EQU Radial stress: S.sub.R =(1.33te+1)M.sub.o /Lt.sup.2 B (4) EQU Tangential stress: S.sub.T =YM.sub.o /t.sup.2 B-ZS.sub.R ( 5)
While it is not the purpose of this background to set forth a dissertation of flange design, such technology being known by the ordinarily skilled person in this field, it may be seen that the flange stress, in all instances, is linearly proportional to the flange moment "M.sub.o ". For gasket seating, the total flange moment "M.sub.o " is based on the flange design bolt load, which is opposed only by the gasket load in which case: EQU W(C-G)/2 (6)
For high pressure fluid carrying systems, it is heretofore been thought necessary to use gaskets which have a gasket factor "m" and a minimum design seating stress "y" that were relatively high. As is well-known, the system operating pressure equals "y/m". Accordingly, for example, a spiral wound stainless or monel, asbestos filled gasket material has a seating stress of approximately 10000 PSI with a gasket factor of 3.00. Accordingly, such a gasket can handle a system pressure up to approximately 3330 PSI when the minimum design seating stress is applied. In standard high pressure applications, it is typical to use a ring-type gasket having a seating stress value of 18000 to 26000 PSI with a gasket factor of 5.50 to 6.50. Accordingly, these systems can handle system pressures in the 3000 to 4000 PSI range when the minimum design seating stress is applied. Importantly, however, if such gasket materials are placed near the flange bore (annulus), the operating bolt load condition "W.sub.m1 " is minimized. However, such placement increases the flange moment "M.sub.o " since such positioning is furthest away from the bolt holes of the flange. Similarly, placing the gasket nearer to the bolt holes, while reducing the flange moment and thus flange stresses, increases the operating bolt load stress "W.sub.m1 ". Accordingly, in order to allow for greater bolt loads to seat such high pressure gaskets, flange design manufacturers are required to increase the flange dimensions and weight. This typically increases the cost of such flanges dramatically which can concomitantly increase the cost of high pressure flow line systems. Accordingly, there is a substantial and long-felt need for a method of sealing flange joints for high pressure systems without the need for costly flange joints in a manner that still meets ASME/ANSI B16.5 specifications with no ASME Boiler and Pressure Vessel Code deviations.